EP2016349B1 - Heat pump - Google Patents

Heat pump Download PDF

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Publication number
EP2016349B1
EP2016349B1 EP06724016A EP06724016A EP2016349B1 EP 2016349 B1 EP2016349 B1 EP 2016349B1 EP 06724016 A EP06724016 A EP 06724016A EP 06724016 A EP06724016 A EP 06724016A EP 2016349 B1 EP2016349 B1 EP 2016349B1
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EP
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Prior art keywords
water
working
pressure
evaporator
drain
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EP06724016A
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German (de)
French (fr)
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EP2016349A1 (en
Inventor
Holger Sedlak
Oliver Kniffler
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Efficient Energy GmbH
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Efficient Energy GmbH
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Priority to DE202006005461U priority Critical patent/DE202006005461U1/en
Priority to PCT/EP2006/003061 priority patent/WO2007118482A1/en
Priority claimed from EP11158793.7A external-priority patent/EP2343489B1/en
Publication of EP2016349A1 publication Critical patent/EP2016349A1/en
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F24HEATING; RANGES; VENTILATING
    • F24DDOMESTIC- OR SPACE-HEATING SYSTEMS, e.g. CENTRAL HEATING SYSTEMS; DOMESTIC HOT-WATER SUPPLY SYSTEMS; ELEMENTS OR COMPONENTS THEREFOR
    • F24D15/00Other domestic- or space-heating systems
    • F24D15/04Other domestic- or space-heating systems using heat pumps
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04DNON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00Details, component parts, or accessories
    • F04D29/26Rotors specially for elastic fluids
    • F04D29/28Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
    • F04D29/284Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps for compressors
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04DNON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00Details, component parts, or accessories
    • F04D29/26Rotors specially for elastic fluids
    • F04D29/28Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
    • F04D29/30Vanes
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B1/00Compression machines, plant, or systems with non-reversible cycle
    • F25B1/04Compression machines, plant, or systems with non-reversible cycle with compressor of rotary type
    • F25B1/053Compression machines, plant, or systems with non-reversible cycle with compressor of rotary type of turbine type
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B1/00Compression machines, plant, or systems with non-reversible cycle
    • F25B1/10Compression machines, plant, or systems with non-reversible cycle with multi-stage compression
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B27/00Machines, plant, or systems, using particular sources of energy
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B30/00Heat pumps
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B30/00Heat pumps
    • F25B30/02Heat pumps of the compression type
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B39/00Evaporators; Condensers
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B39/00Evaporators; Condensers
    • F25B39/02Evaporators
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F24HEATING; RANGES; VENTILATING
    • F24DDOMESTIC- OR SPACE-HEATING SYSTEMS, e.g. CENTRAL HEATING SYSTEMS; DOMESTIC HOT-WATER SUPPLY SYSTEMS; ELEMENTS OR COMPONENTS THEREFOR
    • F24D2200/00Heat sources or energy sources
    • F24D2200/11Geothermal energy
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B30/00Heat pumps
    • F25B30/06Heat pumps characterised by the source of low potential heat
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B9/00Compression machines, plant, or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002Compression machines, plant, or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F28HEAT EXCHANGE IN GENERAL
    • F28DHEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA DO NOT COME INTO DIRECT CONTACT
    • F28D21/00Heat-exchange apparatus not covered by any of the groups F28D1/00 - F28D20/00
    • F28D2021/0019Other heat exchangers for particular applications; Heat exchange systems not otherwise provided for
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F28HEAT EXCHANGE IN GENERAL
    • F28DHEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA DO NOT COME INTO DIRECT CONTACT
    • F28D21/00Heat-exchange apparatus not covered by any of the groups F28D1/00 - F28D20/00
    • F28D2021/0019Other heat exchangers for particular applications; Heat exchange systems not otherwise provided for
    • F28D2021/0068Other heat exchangers for particular applications; Heat exchange systems not otherwise provided for for refrigerant cycles
    • F28D2021/007Condensers
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F28HEAT EXCHANGE IN GENERAL
    • F28DHEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA DO NOT COME INTO DIRECT CONTACT
    • F28D21/00Heat-exchange apparatus not covered by any of the groups F28D1/00 - F28D20/00
    • F28D2021/0019Other heat exchangers for particular applications; Heat exchange systems not otherwise provided for
    • F28D2021/0068Other heat exchangers for particular applications; Heat exchange systems not otherwise provided for for refrigerant cycles
    • F28D2021/0071Evaporators
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F28HEAT EXCHANGE IN GENERAL
    • F28FDETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
    • F28F13/00Arrangements for modifying heat-transfer, e.g. increasing, decreasing
    • F28F13/06Arrangements for modifying heat-transfer, e.g. increasing, decreasing by affecting the pattern of flow of the heat-exchange media
    • F28F13/08Arrangements for modifying heat-transfer, e.g. increasing, decreasing by affecting the pattern of flow of the heat-exchange media by varying the cross-section of the flow channels
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F28HEAT EXCHANGE IN GENERAL
    • F28FDETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
    • F28F21/00Constructions of heat-exchange apparatus characterised by the selection of particular materials
    • F28F21/06Constructions of heat-exchange apparatus characterised by the selection of particular materials of plastics material
    • YGENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02ATECHNOLOGIES FOR ADAPTATION TO CLIMATE CHANGE
    • Y02A40/00Adaptation technologies in agriculture, forestry, livestock or agroalimentary production
    • Y02A40/90Adaptation technologies in agriculture, forestry, livestock or agroalimentary production in food processing or handling
    • Y02A40/96Adaptation technologies in agriculture, forestry, livestock or agroalimentary production in food processing or handling relating to food management or storing
    • Y02A40/963Off-grid food refrigeration
    • Y02A40/965Devices using heat pumps
    • YGENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02BCLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO BUILDINGS, e.g. HOUSING, HOUSE APPLIANCES OR RELATED END-USER APPLICATIONS
    • Y02B10/00Integration of renewable energy sources in buildings
    • Y02B10/40Geothermal heat-pumps

Abstract

A heat pump comprises an evaporator (10) for the evaporation of water as a working liquid, to generate a working vapor, whereby the evaporation takes place at an evaporation pressure of less than 20 hPa. The working vapor is condensed by a compressor (16) to a working pressure of at least 25 hPa, then to be liquefied in a liquefier (18) by direct contact with liquefier water. The heat pump is preferably an open system, in which water present in the environment in the form of ground water, sea water, river water, lake water or brine is evaporated, and reliquefied water is fed to the evaporator, to the ground or to a purification plant.

Description

  • The present invention relates to heat pumps and more particularly to heat pumps that can be used for building heating and especially for heating buildings of smaller building units, such as single-family homes, semi-detached houses or terraced houses.
  • Fig. 8 shows a known heat pump, as in " Technical Thermodynamics ", Theoretical Foundations and Practical Applications, 14th revised edition, Hanser Verlag, 2005, pages 278 - 279 , is described. The heat pump includes a closed circuit in which a working fluid such as R 134a circulates. About a first heat exchanger 80 and the evaporator so much heat is removed from the soil or groundwater that the working fluid evaporates. The now high-energy working fluid is sucked from the compressor via the suction line. In the compressor 81, it is compressed, thereby increasing pressure and temperature. This compression is performed by a reciprocating compressor. The compressed and under high temperature working fluid now passes into the second heat exchanger 82, the condenser. In the condenser so much heat is removed from the working fluid from the heating or hot water circuit that the refrigerant is liquefied under high pressure and high temperature. In throttling or expansion member 83, the working fluid is expanded, ie the working fluid is released. Here, the pressure and temperature are reduced to such an extent that the working fluid is again able to absorb energy from the soil or groundwater in the evaporator again. The cycle is now closed and starts again.
  • As can be seen from this, the working medium serves as an energy transporter to pick up heat from the ground or groundwater and deliver it to the heating circuit in the condenser. In this litigation, the second law is the Thermodynamics fulfilled, in which it is said that heat or energy from "self" can only be transferred from the higher temperature level to the lower temperature level, and that conversely this can only happen by external energy supply, here the drive work of the compressor.
  • Fig. 7 shows a typical h, log p diagram (h is the enthalpy, p the pressure of a substance). Between point 4 and point 1 in the diagram of Fig. 7 an isobaric evaporation of the working fluid takes place at low values for the pressure and the temperature (p 1 , T 1 ). Here, the heat Q 81 is supplied.
  • Between point 1 and point 2, ideally, a reversible compression of the working medium vapor in an adiabatic compressor to the pressure p 2 takes place. The temperature rises to T 2 . It is here to perform a compression work.
  • Then, at high pressure p 2 , an isobaric cooling of the working medium vapor is first carried out from 2 to 2 '. The overheating is reduced. Subsequently, a liquefaction of the working fluid takes place. Overall, the heat Q 25 can be dissipated.
  • In the throttle 83, the adiabatic throttling of the working fluid then takes place from the high pressure p 2 to the low pressure p 1 . In this case, a part of the liquid working fluid evaporates and the temperature decreases to the evaporation temperature T 1 . In the h, log p diagram, the energies and characteristics of this process can be calculated by means of enthalpies and illustrated as in Fig. 7 is shown.
  • The working fluid of the heat pump thus takes in the evaporator heat from the environment, ie air, water, sewage or soil, on. The condenser serves as a heat exchanger for heating a heating medium. The temperature T 1 is slightly below the ambient temperature, the temperature T 2 considerably, the temperature T 2 'slightly above the required heating temperature. The higher the required temperature difference, the more work the compressor must apply. It is therefore desirable to keep the temperature increase as small as possible.
  • Referring to Fig. 7 Thus, in the ideal case, a compression of the working substance vapors along the curve for the entropy s = constant up to point 2 is performed. From here to point 3 the working fluid liquefies. The length of the section 2-3 represents the useful heat Q. From the point 3 to the point 4, the relaxation takes place and from point 4 to point 1, the evaporation of the working substance, the distance 4-1 reproduces the heat extracted from the heat source. In contrast to the T, s diagram, the amounts of heat and work can be taken as stretches in the h, log p diagram. Pressure losses in valves, the pressure and suction lines, the compressor, etc. deform the ideal course of the cycle in the h, log p diagram and reduce the effectiveness of the entire process.
  • In piston compressors, the aspirated working vapor initially has a lower temperature than the cylinder wall of the compressor and absorbs heat from it. As the compression progresses, the temperature of the working-material vapor finally increases above that of the cylinder wall, so that the working-substance vapor gives off heat to the cylinder wall. Then, when the piston again sucks in steam and compressed, the temperature of the piston wall is initially again fallen below and then exceeded, resulting in permanent losses. Furthermore, overheating of the aspirated working fluid will be necessary and necessary so that the compressor does not suck liquid agent. A disadvantage is in particular the heat exchange with the oil circuit of the reciprocating compressor, which is indispensable for lubrication.
  • Occurring irreversibilities, such as heat losses during compression, pressure losses in the valves and flow losses in the pressure line for liquefying and in the condenser increase the entropy, ie the heat that can not be recovered. Furthermore, the temperature T 2 is also above the liquefaction temperature. Such an "overheating enthalpy" is undesirable, especially because the high temperatures occurring accelerate the aging of the compressor and in particular of the lubricating oil in a piston compressor. Also, the effectiveness of the process is reduced.
  • The liquefied low-temperature working fluid at the condenser outlet would have to be relieved by an ideal cycle through an engine, such as a turbine, to utilize the excess of energy that existed against the temperature and pre-compression conditions. For reasons of this required large expense, this measure is omitted and the pressure of the working fluid is abruptly reduced by the throttle 83 to the low pressure and the low temperature. The enthalpy of the working substance remains approximately the same. Due to the sudden reduction in pressure, the working fluid must partially evaporate in order to lower its temperature. The necessary heat of evaporation comes from the excess temperature located working fluid, so it is not removed from the heat source. The totality of the relaxation in the choke 83 ( Fig. 8 ) losses are referred to as relaxation losses. These are exergy losses, because heat of temperature T is converted into heat of temperature T 0 . These losses can be reduced if the liquid agent can deliver its heat to a medium of a temperature less than T. This subcooling enthalpy can be exploited by an internal heat exchange, which, however, again requires additional equipment. In principle, the internal heat exchange is limited, because in the compression of the vapors, the superheating temperature T 2 increases, whereby the profits achieved are partially compensated again, and machine and lubricating oil are subjected to thermal stress. Finally, the overheating increases the volume of the steam, which reduces the volumetric heat output. This heat is used only to preheat the vapors of the working fluid flowing to the compressor as far as necessary to ensure that all the droplets contained in the vapor of the working fluid have been converted to steam with certainty. In general, it can be said that the ratio of the enthalpy difference between point 1 and point 4 to the enthalpy difference between point 2 and point 1 of the h, log p diagram is a measure of the efficiency of the heat pump process ,
  • A currently popular tool is R134a, which has CF 3 -CH 2 F as its chemical formula. This is a working fluid that is no longer harmful to the ozone layer, but has a 1000 times greater effect than carbon dioxide in terms of the greenhouse effect. However, the working fluid R134a is popular because it has a relatively large enthalpy difference of about 150 kJ / kg.
  • Although this working fluid is no longer an "ozone killer", there are still considerable demands on the integrity of the heat pump cycle, such that no working fluid molecules escape from this closed circuit, since these would cause quite considerable damage due to the greenhouse effect. This encapsulation causes significant additional costs when building a heat pump.
  • It is also expected that until the implementation of the next stage of the Kyoto Protocol due to the greenhouse effect, R134a will be banned until 2015, which has been done earlier with much more harmful means.
  • A disadvantage of existing heat pumps is therefore in addition to the fact of the harmful working fluid and the fact that due to the many losses in the heat pump cycle, the efficiency of the heat pump is typically not more than a factor of 3. In other words, about 2 times the energy used for the compressor can be taken from the heat source, such as groundwater or soil. Considering now heat pumps in which the compressor is driven by electric current, and at the same time taking into account that the efficiency in power generation is perhaps equal to 40%, so turns out that - in terms of the overall energy balance - a heat pump is doubtful in terms of utility. Based on the primary energy source, 120% = 3 40% of heating energy is provided. After all, a conventional heating system with a burner achieves efficiencies of 90 - 95%, ie with a high technical and thus financial expense only an improvement of 25 - 30% is achieved.
  • Better systems use primary energy to drive the compressor. So gas or oil is burned to create the compressor power with the energy released by the combustion. An advantage of this solution is that the energy balance is actually more positive. This has as a reason that only about 30% of the primary energy carrier can be obtained as drive energy, but for the waste heat of then about 70% can be used for heating. The heating energy provided is then 160% = 3 30% + 70% of the primary energy source. A disadvantage of this solution, however, is that, a household, although he has no more classical heating, still requires an internal combustion engine and a fuel storage. The cost of engine and fuel storage are still added to the cost of the heat pump, which is a high-closed circuit due to the climate-damaging coolant.
  • All of these things have meant that heat pumps can only compete to a certain extent in competition with other types of heating.
  • The DE 44 31 887 A1 discloses a heat pump system according to the preamble of claim 1 with a propeller-like rotary member consisting of a frusto-conical hub and a plurality of curved blades. Further, a mechanical steam compression heat pump system is provided which includes a pair of centrifugal compressors.
  • The DE 10 2004 001 927 A1 discloses a method for heat recovery by means of a heat pump, in which a screw compressor with oil cooling and oil injection and water are used as refrigerant.
  • The U.S. Patent No. 4,003,213 discloses a heat pump system according to the preamble of claim 1 which is suitable for the production of drinking water, and in which the thermal energy is generated from heat released by the freezing of water under vacuum at the triple point.
  • The US 6,397,621 B1 discloses a heat pump installation according to the preamble of claim 1 having a cooling function in which the thermodynamic fluid used in the cycle is water. Dynamic compression in two separate compression stages is used, with the compression stages interconnected by at least one zone preventing overheating.
  • The object of the present invention is to provide a more efficient heat pump concept.
  • This object is achieved by a heat pump according to claim 1, or a method for pumping heat according to claim 9.
  • The present invention is based on the recognition that it is necessary to move away from climate-damaging working materials, and that instead normal water is an optimal working medium. In addition, water has a much larger ratio of enthalpy differences than the currently widely used R134a. The enthalpy difference, which is crucial to the effectiveness of the heat pump process, is about 2500 kJ / kg for water, which is about 16 times the useful enthalpy difference of R134a. By contrast, the compressor enthalpy to be used is only 4 to 6 times greater, depending on the operating point.
  • In addition, water is not climate damaging, so neither an ozone killer nor a greenhouse effect intensifier. This allows heat pumps to be made much simpler, since the demands on the integrity of the circuit are not high. Instead, it is even preferred to go completely away from the closed process and instead to make an open process in which the groundwater or water, which is the external heat source, is directly evaporated.
  • According to the invention, the evaporator is designed such that it has an evaporation space in which the evaporation pressure is less than 20 hPa (hectopascal), so that the water evaporates at temperatures below 18 ° C and preferably below 15 ° C. Typical groundwater has temperatures in the northern hemisphere between 8 and 12 ° C, which requires pressures of less than 20 hPa, so that the groundwater evaporates in order to be able to achieve a lowering of the temperature of the groundwater and thus a heat extraction by the groundwater evaporation a building heating, such as a floor heating can be operated.
  • Water is further advantageous in that water vapor occupies a very large volume, and thus to the compaction ten of the steam no longer on a displacement machine such as a piston pump or something similar must be resorted to, but a high-performance compressor in the form of a turbomachine such as a centrifugal compressor can be used, which is well manageable in the art and is inexpensive to manufacture, since he exists in large numbers and is used for example as a small turbine or as a turbocompressor in cars so far.
  • A prominent representative of the class of turbomachinery in comparison to displacement machines is the radial compressor, for example in the form of a centrifugal compressor with radial wheel.
  • The radial compressor or turbomachine must achieve at least one compression that the outlet pressure from the centrifugal compressor is higher by 5 hPa than the inlet pressure into the centrifugal compressor. Preferably, however, a densification in a ratio greater than 1: 2 and even greater than 1: 3 will be.
  • Turbomachines also have the advantage compared to typically used in closed circuits Kolbenverdichtern that the compressor losses due to the existing temperature gradient in the turbomachine compared to a displacement machine (reciprocating compressor), in which such a stationary temperature gradient does not exist, are greatly reduced. It is particularly advantageous that an oil circuit completely eliminated.
  • Furthermore, multi-stage turbomachines are particularly preferred in order to achieve the relatively high compression, which, in order to achieve a sufficient flow temperature of a heater for cold winter days, should have a factor of 8 to 10.
  • In a preferred embodiment, a completely open circuit is used in which groundwater is brought to the low pressure. A preferred embodiment for generating a pressure below 20 hPa for groundwater is the simple use of a riser, which in a pressure-tight evaporation chamber opens. If the riser crosses a height between 9 and 10 m, then the required low pressure is present in the evaporation space, at which the groundwater evaporates at a temperature between 7 and 12 ° C. After typical buildings are at least 6 to 8 meters high, and since in many regions the groundwater already at 2 to 4 m below the surface, the installation of such a pipe does not lead to significant additional effort, as only slightly lower than for the house foundation must be dug, and because typical building heights are readily so high that the riser or the evaporation chamber does not protrude beyond the building.
  • For applications where only a shorter riser is possible, the length of the riser can be readily reduced by a pump / turbine combination, due to the fact that the turbine is designed for high to low pressure conversion and the pump for conversion used by the low pressure on the high pressure, only a little extra work needed from the outside.
  • This primary heat exchanger losses are eliminated because no primary heat exchanger is used, but the evaporated groundwater is used directly as working steam or working fluid.
  • In a preferred embodiment, no heat exchanger is used in the condenser. Instead, the water vapor heated due to its compression is introduced directly into the heating water in a condenser, so that liquefaction of the water vapor takes place within the water, such that secondary heat exchanger losses are also eliminated.
  • The water evaporator fluid machine condenser combination according to the invention thus enables efficiencies of at least factor 6 in comparison to conventional heat pumps. Thus, at least 5 times the electrical energy expended in the compression can be extracted from the groundwater be so that even if the turbomachine is operated with electric current, a heating energy of 240% = 6 40% based on the primary energy carrier is provided. This is at least a doubling of efficiency compared to the prior art, or halved compared to energy costs. This applies in particular to climate-relevant emissions of carbon dioxide.
  • Preferred embodiments of the present invention will be explained below in detail with reference to the accompanying drawings. Show it:
  • Fig. 1a
    a schematic block diagram of the inventive heat pump;
    Fig. 1b
    a table illustrating various pressures and the evaporation temperatures associated with these pressures;
    Fig. 2
    a block diagram of a preferred exemplary embodiment of the heat pump according to the invention, which is operated with groundwater, seawater, river water, seawater or brine;
    Fig. 3a
    an alternative embodiment of the condenser of Fig. 2 ;
    Fig. 3b
    an alternative embodiment of the condenser with reduced return in off mode;
    Fig. 3c
    a schematic representation of the condenser with a gas separator;
    Fig. 4a
    a preferred implementation of the evaporator of Fig. 2 ;
    Fig. 4b
    an alternative embodiment of the evaporator with the use of the condenser drain as Siedeunter- support;
    Fig. 4c
    an alternative embodiment of the evaporator with a heat exchanger for use of groundwater for Siedeunterstützung;
    Fig. 4d
    an alternative embodiment of the evaporator with feed from the side and drain in the middle;
    Fig. 4e
    a schematic representation of the expander with indication of preferred dimensions;
    Fig. 5a
    an alternative implementation of the evaporator to reduce the height of the riser;
    Fig. 5b
    an implementation of an alternative implementation of a connection of a heating line to the condenser with a turbine / pump combination;
    Fig. 6a
    a schematic representation of the compressor performed by several sequentially arranged flow machines;
    Fig. 6b
    a schematic representation of the setting of the rotational speeds of two cascaded turbomachines in dependence on the target temperature;
    Fig. 6c
    a schematic plan view of a radial impeller of a turbomachine according to a preferred embodiment of the present invention;
    Fig. 6d
    a schematic cross-sectional view with a merely schematic representation of the Radialrad- blades to illustrate the different extent of the blades with respect to the radius of the radial wheel;
    Fig. 7
    an exemplary h, log p diagram; and
    Fig. 8
    a well-known heat pump, the left-handed cycle of Fig. 7 performs.
  • Fig. 1a shows a heat pump according to the invention, which initially has a water evaporator 10 for evaporating water as a working fluid to the output side to generate a steam in a working steam line 12. The evaporator comprises an evaporation space (in Fig. 1a not shown) and is designed to produce an evaporation pressure of less than 20 hPa in the evaporation space, so that the water evaporates at temperatures below 15 ° C in the evaporation space. The water is preferably groundwater, in the ground free or in collector pipes circulating brine, so water with a certain salinity, river water, seawater or seawater. According to the invention, all types of water, ie, calcareous water, lime-free water, saline water, or salt-free water, are usefully used. This is because all types of water, that is, all of these "hydrogens", have the favorable water property, namely that water, also known as "R 718", is an enthalpy difference ratio useful for the heat pump process of 6 has, which is more than 2 times the typical usable enthalpy difference ratio of z. B. R134a corresponds.
  • The water vapor is supplied through the suction line 12 to a compressor / condenser system 14, which is a turbomachine such. B. has a centrifugal compressor, for example in the form of a turbocompressor, in Fig. 1a denoted by 16. The turbomachine is designed to compress the working steam to a vapor pressure at least greater than 25 hPa. 25 hPa corresponds to a liquefaction temperature of about 22 ° C, which can already be a sufficient heating flow temperature of a floor heating, at least on relatively warm days. In order to generate higher flow temperatures, pressures greater than 30 hPa can be generated with the turbomachine 16, wherein a pressure of 30 hPa has a liquefaction temperature of 24 ° C, a pressure of 60 hPa has a liquefaction temperature of 36 ° C, and a pressure of 100 hPa corresponds to a liquefaction temperature of 45 ° C. Underfloor heating systems are designed to heat sufficiently with a flow temperature of 45 ° C, even on very cold days.
  • The turbomachine is coupled to a condenser 18, which is designed to liquefy the compressed working steam. By liquefying the energy contained in the working steam is supplied to the condenser 18, to then be supplied via the flow 20a a heating system. About the return 20b, the working fluid flows back into the condenser.
  • According to the invention, it is preferable to extract from the high-energy steam directly through the colder heating water the heat (energy) which is taken up by the heating water so that it heats up. The steam is so much energy withdrawn that this is liquefied and also participates in the heating circuit.
  • Thus, a material entry into the condenser or the heating system takes place, which is regulated by a drain 22, such that the condenser has a water level in its condenser, which remains despite the constant supply of water vapor and thus condensate always below a maximum level.
  • As has already been stated, it is preferred to take an open circuit, ie to evaporate the water, which is the heat source, directly without a heat exchanger. Alternatively, however, the water to be evaporated could first be heated by a heat exchanger from an external heat source. However, it should be remembered that this heat exchanger again means losses and equipment expense.
    Moreover, in order to avoid losses for the second heat exchanger, which has hitherto necessarily been present on the condenser side, it is also preferred to use the medium directly there as well, that is to say to a house with underfloor heating It is thought that the water coming from the evaporator should be circulated directly in the underfloor heating.
  • Alternatively, however, on the condenser side, a heat exchanger can be arranged, which is fed with the flow 20a and having the return 20b, said heat exchanger cools the water in the condenser and thus heats a separate underfloor heating fluid, which will typically be water.
  • Due to the fact that water is used as the working medium, and due to the fact that only the evaporated portion of the groundwater is fed into the turbomachine, the purity of the water does not matter. The turbomachine, as well as the condenser and possibly directly coupled underfloor heating always supplied with distilled water, so that the system has a reduced maintenance compared to today's systems. In other words, the system is self-cleaning, since the system is always fed only distilled water and the water in the drain 22 is thus not polluted.
  • In addition, it should be noted that turbomachines have the properties that they - similar to an aircraft turbine - the compressed medium not with problematic substances, such as oil, in connection. Instead, the water vapor is compressed only by the turbine or the turbo compressor, but not associated with oil or other purity-impairing medium and thus contaminated.
  • The distilled water discharged through the drain can thus - if no other regulations stand in the way - be easily returned to the groundwater. Alternatively, however, it may also be z. B. in the garden or in an open space to be seeped, or it can be supplied via the channel, if regulations dictate - a sewage treatment plant.
  • The combination according to the invention of water as a working medium with the two times better usable enthalpy difference ratio compared to R134a and due to the reduced requirements for the closed system (rather an open system is preferred), and because of the use the turbomachine, through which the required compression factors are achieved efficiently and without purity impairments, an efficient and environmentally neutral heat pump process is created, which is even more efficient if the liquefier in the liquefier directly liquefied, since then no heat exchanger is needed in the entire heat pump process ,
  • In addition, all associated with the piston compression losses fall away. In addition, the very low water losses which would otherwise be incurred in the throttling can be used to improve the evaporation process, since the effluent water at the effluent temperature, which will typically be higher than the groundwater temperature, will be used to advantage Evaporator by means of a structuring 206 of a drain pipe 204, as shown in Fig. 4a is still to trigger a bubble evaporation, so that the evaporation efficiency is increased.
  • Hereinafter, referring to Fig. 2 a preferred embodiment of the present invention explained in detail. The water evaporator comprises a vaporization chamber 100 and a riser 102 in which groundwater from a groundwater reservoir 104 moves upwardly in the direction of an arrow 106 into the vaporization space 100. The riser 102 opens into an expander 108, which is designed to widen the relatively narrow tube cross-section in order to create the largest possible evaporation surface. The expander 108 will be funnel-shaped, that is to say in the form of a paraboloid of revolution of any shape. It can have round or angular transitions. The decisive factor is that the diameter directed into the evaporation chamber 100 or the surface facing the evaporation chamber 100 is greater than the cross-sectional area of the riser, in order to improve the evaporation process. If it is assumed that about 1 l per second flows through the riser up into the evaporation chamber, about 4 ml per second are evaporated in the evaporator at a heating power of about 10 kW. The remainder, cooled by about 2.5 ° C, passes over the expander 108 and lands in a collection sump 110 in the vaporization chamber. The collecting sump 110 has a drain 112, in which the amount of 1 1 per second less the evaporated 4 ml per second is discharged again, preferably back into the groundwater reservoir 104. For this purpose, preferably a pump 114 and a valve for Overflow control provided. It should be noted that nothing has to be pumped actively here, since due to gravity, when the pump or the valve 114 is opened, water flows from the evaporator catch basin 110 via a return pipe 113 down into the groundwater reservoir. The pump or the valve 114 thus ensure that the water level in the catch basin does not rise too high or that no water vapor penetrates into the drain pipe 112 or that the evaporation chamber is reliably decoupled from the situation at the "lower" end of the return pipe 113.
  • The riser is arranged in a riser 116, which is filled by a preferably provided pump 118 with water. The levels in 116 and 108 are interconnected according to the communicating tube principle, with gravity and the different pressures in 116 and 108 providing water transport from 116 to 108. The water level in the riser tank 116 is preferably arranged so that even at different air pressures, the level never falls below the inlet of the riser 102, so that the ingress of air is avoided.
  • Preferably, the evaporator 10 includes a gas separator configured to receive at least a portion, e.g. To remove at least 50% of a gas dissolved in the water to be evaporated, from the water to be evaporated, so that the removed part of the gas is not sucked through the evaporation space of the compressor. Preferably, the gas separator is arranged to a remote part of the gas not supplied evaporated water, so that the gas is transported away from the unevaporated water. Dissolved gases may include oxygen, carbon dioxide, nitrogen, etc. These gases usually evaporate at a higher pressure than water so that the gas separator can be arranged below the expander 108, so that oxygen vaporized in the gas separator etc. escapes from the water which is not yet evaporating and is preferably fed into the return line 113. The feed is preferably carried out at the point of the return line 113, at which the pressure is so low that the gas is taken back into the groundwater by the returning water. Alternatively, however, the separated gas can also be collected and disposed of at certain intervals or continuously vented, that is delivered to the atmosphere.
  • Typically, the groundwater, seawater, river water, seawater, brine or any other naturally occurring aqueous solution will have a temperature between 8 ° C and 12 ° C. By lowering the temperature of 1 liter of water by 1 ° C, a power of 4.2 kW can be generated. If the water is cooled by 2.5 ° C, a power of 10.5 kW is generated. Preferably, the riser is flowed through by a stream of water with a current in dependence on the heating power, in the example one liter per second.
  • When the heat pump is operating at a relatively high load, the evaporator will vaporize about 6 ml per second, which corresponds to a vapor volume of about 1.2 cubic meters per second. Depending on the required heating water temperature, the turbomachine is controlled with regard to its compaction performance. If a heating flow temperature of 45 ° C is desired, which is by far sufficient even for extremely cold days, then the turbomachine must increase the steam produced at perhaps 10 hPa to a pressure of 100 hPa. In contrast, a flow temperature of z. B. 25 ° for underfloor heating, so only by a factor of 3 must be compressed by the turbomachine.
  • The power generated is therefore determined by the compressor power, so on the one hand by the compression factor, ie how much compressed the compressor, and on the other by the volume flow generated by the compressor. As the volumetric flow increases, the evaporator must vaporize more, with the pump 118 conveying more groundwater into the riser basin 116, so that more groundwater is supplied to the vaporization chamber. If, on the other hand, the turbomachine delivers a lower compression factor, less groundwater flows from the bottom to the top.
  • It should be noted, however, that it is preferred to control the flow of groundwater through the pump 118. According to the principle of communicating tubes, the level in the container 116 or the delivery rate of the pump 118 defines the flow through the riser. Thus, an increase in efficiency of the system can be achieved because the control of the flow is decoupled from the suction power of the turbomachine.
  • No pump is needed to pump the groundwater from below into the vaporization chamber 100. Instead, this happens from "self". This automatic upgrade to the evacuated evaporation chamber also helps in that the negative pressure of 20 hPa is readily achievable. For this purpose, no evacuation pumps or something similar needed. Instead, only a riser with a height greater than 9 m is needed. Then a purely passive vacuum generation is achieved. However, the necessary negative pressure can also be generated with a much shorter riser, if z. B. the implementation of Fig. 5a is used. In Fig. 5a is shown a much shortened "riser". The implementation of the high pressure to the negative pressure is effected via a turbine 150, wherein the turbine extracts energy from the working medium. At the same time the negative pressure on the return side is brought back into the high pressure, the energy required for this is supplied by a pump 152. The pump 152 and the turbine 150 are coupled together via a power coupling 154 so that the turbine drives the pump, with the energy that the turbine has removed from the medium. A motor 156 is only needed to compensate for the losses that the system has of course, and to achieve the circulation, so as to move a system from its rest position into the in Fig. 5a to bring shown dynamic mode.
  • In the preferred embodiment, the turbomachine is configured as a rotary compressor with a rotatable wheel, wherein the wheel may be a low-speed radial, medium-radial, semi-axial, or propeller, as known in the art. Radial compressors are described in "Turbomachines", C. Pfleiderer, H. Petermann, Springer-Verlag, 2005, pages 82 and 83. Such radial compressors thus include as a rotatable wheel the so-called. Center runner whose form depends on the individual requirements. In general, any flow machines can be used, as they are known as turbo compressors, fans, blowers or turbocompressors.
  • In the preferred embodiment of the present invention, the radial compressor 16 is designed as a plurality of independent turbomachines, which can be controlled independently of each other at least in terms of their speed, so that two turbomachines can have different speeds. Such an implementation is in Fig. 6a represented, in which the compressor is designed as a cascade of n turbomachines. At any point after the first turbomachine one or more heat exchangers, for example, for hot water heating, which are denoted by 170, is preferably provided. These heat exchangers are designed to cool the gas which has been heated (and compressed) by a preceding turbomachine 172. Here, the superheat enthalpy is meaningfully used to increase the efficiency of the entire compression process. The cooled gas is then further compressed with one or more downstream compressors or fed directly to the condenser. It is removed heat from the compressed water vapor, so that z. As process water to higher temperatures than z. B. 40 ° C to heat. However, this does not reduce the overall efficiency of the heat pump, but even increases it, since two consecutively connected gas cooling turbines with a longer service life achieve the required gas pressure in the condenser due to the reduced thermal stress and less energy than if a single turbomachine without gas cooling would be available.
  • The cascaded independently operated turbomachines are preferably controlled by a controller 250, the input side receives a desired temperature in the heating circuit and possibly also an actual temperature in the heating circuit. Depending on the desired setpoint temperature, the rotational speed of a turbomachine arranged earlier in the cascade, which is denoted by n 1 by way of example, and the rotational speed n 2 of a turbomachine, which is arranged later in the cascade, are changed as indicated by Fig. 6b is shown. If a higher set temperature is input to the controller 250, both speeds are increased. However, the speed of the previously arranged turbomachine, with n 1 in Fig. 6b is increased with a smaller gradient than the rotational speed n 2 of a later arranged in the cascade turbo machine. As a result, when the ratio n 2 / n 1 of the two speeds is plotted, the graph of Fig. 6b gives a straight line with a positive slope.
  • The intersection between the individually applied rotational speeds n 1 and n 2 can take place at any desired point, that is to say at any desired temperature and, if appropriate, can not take place. In general, however, it is preferable to lift a turbomachine arranged closer to the condenser in the cascade more strongly in terms of its rotational speed than a turbomachine arranged earlier in the cascade, if a higher desired temperature is desired.
  • The reason for this is that the later arranged in the cascade turbomachine already compressed gas, the has been compressed by an earlier arranged in the cascade turbomachine, must process further. Further, this ensures that the blade angle of blades of a radial wheel, as also referring to Fig. 6c and 6d is always as good as possible with respect to the speed of the gas to be compressed. Thus, the adjustment of the blade angle is only in the optimization of possible low-turbulence compression of the incoming gas. The other parameters of the angle adjustment such as gas flow rate and compression ratio, which would otherwise have made a technical compromise in the choice of blade angle and thus only at a desired temperature optimum efficiency are brought according to the invention by the independent speed control to its optimum operating point and therefore have no influence more on the choice of blade angle. This results in spite of a fixed blade angle always optimal efficiency.
  • In view of this, it is further preferred that a turbomachine arranged more in the cascade in the direction of the condenser has a direction of rotation of the radial wheel which is opposite to the direction of rotation of a radial wheel previously arranged in the cascade. Thus, a nearly optimal entry angle of the blades of both axial wheels can be achieved in the gas flow, such that a favorable efficiency of the turbomachine cascade occurs not only in a small target temperature range, but in a much larger target temperature range between 20 and 50 degrees, which is an optimal range for typical heating applications. The speed control according to the invention and, where appropriate, the use of counter-rotating axial gears thus provides an optimal match between the variable gas flow with changing target temperature on the one hand and the fixed blade angles of the axial wheels on the other.
  • In preferred embodiments of the present invention, at least one or preferably all of the axial gears of all flow machines are made of plastic having a tensile strength above 80 MPa. A preferred plastic this is polyamide 6.6 with inserted carbon fibers. This plastic has the advantage of tensile strength, so that Axialräder the fault machines can be made of this plastic and yet can be operated at high speeds.
  • Axial wheels are preferably used according to the invention, as for example in Fig. 6c at reference numeral 260. Fig. 6c shows a schematic plan view of such a radial wheel, wherein Fig. 6d a schematic cross-sectional view of such a radial wheel shows. A radial wheel, as known in the art, includes a plurality of inwardly outwardly extending vanes 262. The vanes extend from a distance of a central axis 264, designated r w , entirely outward with respect to the axis 264 of the radial wheel. In particular, the radial wheel comprises a base 266 and a lid 268 which is directed to the intake manifold or to an earlier stage compressor. The radial wheel comprises a suction port, designated r 1 , for sucking gas, which gas is then discharged laterally from the radial wheel, as indicated at 270 in FIG Fig. 6d is specified.
  • If Fig. 6c Thus, in the direction of rotation in front of the blade 262, the gas is at a relatively higher velocity, while behind the blade 262 it is at a reduced velocity. However, for high efficiency and high efficiency, it is preferred that the gas be evacuated laterally from the radial wheel at a rate as uniform as possible, ie at 270 in Fig. 6d is ejected. For this purpose, there is a desire to mount the blades 262 as close as possible.
  • However, for technical reasons, an arbitrarily tight attachment of from the inside, ie from the radius r w outwardly extending blades is not possible because then the suction port with the radius r 1 is more and more blocked.
  • In accordance with the invention, it is therefore preferred to provide blades 272, 274 and 276, respectively, which are less long than the blade 262 extend. In particular, the vanes 272 do not extend from r w to the outside, but outwardly from R 1 with respect to the radial wheel, where R 1 is greater than r w . Analogously extend, as in Fig. 6c by way of example only, the vanes 274 extend outwardly from R 2 only, while the vanes 276 extend outwardly only from R 3 , where R 2 is greater than R 1 and R 3 is greater than R 2 .
  • These relationships are in Fig. 6d shown schematically, with a double hatching, for example in the area 278 in Fig. 6d indicates that there are two blades in this area that overlap and are therefore marked by the double-hatched area. Example, the hatching from bottom left to top right in the area 278 denotes a vane 262 w extends from r to by the very outside, while extending from top left to bottom right of the area 278 hatching indicates a vane 272 which merely to extends from R 1 to the outside with respect to the radial wheel.
  • Preferably, at least one blade, which does not extend so far inwardly, is thus arranged between two blades extending deeper inwardly. This means that the intake is not clogged or areas with a smaller radius are not too heavily occupied with blades, while areas with a larger radius are more densely occupied with blades, so that at the exit of the radial wheel, ie where the compressed gas Radial wheel leaves, as homogeneous a velocity distribution of the exiting gas exists. The velocity distribution of the exiting gas is in the preferred radial wheel according to the invention in Fig. 6c particularly homogeneous on the outer circumference, because the distance of blades accelerating the gas and due to the "stacked" arrangement of the blades is substantially smaller than in a case where, for example, only the blades 262 are present, which are from the very inside extend to the very outside and thus inevitably have at the outer end of the radial wheel a very large distance, which is much larger as in the radial wheel according to the invention, as in Fig. 6c is shown.
  • At this point it should be noted that the relatively complex and complicated shape of the radial wheel in Fig. 6c Particularly favorable can be produced with plastic injection molding, which in particular can be easily achieved that all blades, including the blades that do not extend from the very inside to the very outside, so the blades 272, 274, 276 are firmly anchored, as they both with the lid 268 so on the basis 266 of Fig. 6d are connected. The use of plastic in particular with the plastic injection molding technology makes it possible to produce any shapes accurately and inexpensively, which is not readily possible with axial wheels made of metal or only very expensive or possibly even impossible.
  • It should be noted at this point that very high rotational speeds of the radial wheel are preferred, so that the accelerations acting on the blades assume very considerable values. For this reason, it is preferred that, in particular, the shorter blades 272, 274, 276 are fixedly connected not only to the base but also to the cover, such that the radial wheel can readily withstand the accelerations that occur.
  • In this context, it should also be noted that the use of plastic is also favorable due to the superior impact resistance of plastic. So it is not always ruled out that ice crystals or water droplets hit the radial wheel at least the first compressor stage. Due to the high accelerations, very high impact forces arise here, which are easily withstood by plastics with sufficient impact resistance. Furthermore, the liquefaction in the liquefier preferably takes place on the basis of the cavitation principle. Here steam bubbles fall due to this principle in a volume of water in itself. There are also microscopically quite considerable velocities and forces arising over the long term Seen view can lead to material fatigue, which, however, when a plastic is used with a sufficient impact resistance, are easily manageable.
  • The compressed gas discharged from the last compressor 174, that is, the compressed steam, is then supplied to the condenser 18, which may be configured as shown in FIG Fig. 2 is shown, but which is preferably designed as in Fig. 3a is shown. The condenser 18 contains a volume of water 180, and preferably an arbitrarily small volume of steam 182. The condenser 18 is designed to feed the compressed vapor into the water of the volume of water 180, so that where the vapor enters the liquid immediately results in condensation as schematically indicated at 184. For this purpose, it is preferred that the gas supply has a widening region 186, such that the gas is distributed as extensively as possible in the condenser water volume 180. Typically, due to the temperature layers in a water tank above, the highest temperature and below the coolest temperature. Therefore, via a float 188, the heating flow is placed as close as possible to the surface of the water volume 180 to always remove the warmest water from the condenser water volume 180. The heating return is fed down the condenser, so that the steam to be liquefied always comes in contact with the coolest possible water, which due to the circulation using a heating circulation pump 312 again from below towards the steam-water boundary of the expander 186 moves.
  • The embodiment in Fig. 2 in which only a single circulation pump 312 exists, it is sufficient if the condenser is arranged in a building so that the areas to be heated are below the condenser, so that due to the gravity in all the heating pipes is a greater pressure than in the condenser.
  • Fig. 5b on the other hand shows an implementation of a connection of a heating line to the condenser with a turbine / pump combination, if the condenser is to be located at a lower height than the heating line, or if a conventional heater, which requires a higher pressure, to be connected. Thus, if the condenser is to be arranged lower, that is to say below a surface to be heated or the heating line 300, then the pump 312 is designed as a driven pump, as at 312 in FIG Fig. 5b is shown. Further, a turbine 310 is provided in the heater return 20b for driving the pump 312, which is connected via a power coupling 314 to the pump 312. The high pressure then prevails in the heater and the low pressure prevails in the condenser.
  • After due to the constantly introduced into the condenser, the water level in the condenser would increase more and more, the drain 22 is provided above which, so that the water level in the condenser does not change substantially, also z. B. have to drain about 4 ml per second. For this purpose, a drain pump or a drain valve 192 is provided for pressure control, such that without pressure loss, the required amount of z. B. 4 ml per second, ie the amount that is supplied to the steam liquefier with the compressor running, is discharged again. Depending on the implementation, the drain may be introduced into the riser, as shown at 194. After all the pressures between a bar and the present in the evaporation space pressure along the riser 102, it is preferred to feed the outlet 22 at the point 194 in the riser at which approximately the same pressure exists as he after the pump 192 and the Valve 192 is present. Then no work must be done to supply the drain water to the riser again.
  • At the in Fig. 2 embodiment shown is carried out completely without a heat exchanger. The groundwater is thus evaporated, the steam is then liquefied in the condenser, and the liquefied steam is finally pumped through the heater and fed back to the riser. However, after not all of the water flowing through the riser amount of water is evaporated, but always only a (very small) share, is Thus, water that has flowed through the underfloor heating, fed to the groundwater. If such a thing is prohibited by local regulations, although the present invention does not involve any contamination, the process may also be designed to deliver the amount of 4 ml per second corresponding to about 345 liters per day to the channel. This would ensure that no medium that was in a heating system of a building is directly fed back into the groundwater. '
  • However, the return 112 from the evaporator can be fed without problems into the groundwater, since the water returning there was in contact only with the riser and the return line, but did not exceed the "evaporation limit" between the evaporator expander 108 and the outlet to the turbomachine Has.
  • It should be noted that the evaporation space at the in Fig. 2 shown embodiment and the condenser or the vapor space 182 of the condenser must be sealed. As soon as the pressure in the evaporation chamber rises above the mark required for the water pumped through the riser pipe to evaporate, the heat pump process will "stop".
  • The following will be on Fig. 3a Reference is made, which represents a preferred embodiment for the condenser 18. The compressed steam supply line 198 is placed in the condenser so that the steam just below the surface of the condenser water volume 180 can escape into this volume of water. For this purpose, the end of the steam supply line around the circumference of the tube arranged around nozzles, through which the steam can escape into the water. To ensure the best possible mixing occurs, so that the steam comes into contact with the coldest possible water in order to liquify as quickly and efficiently as possible, an expander 200 is provided. This expansion is located in the condenser water volume 180. It has at its narrow point a circulating pump 202, which is adapted to suck cold water at the bottom of the condenser and through the expander in to shift an upward broadening flow. This is intended to bring as large amounts of the entering into the condenser water 180 steam with cold water as possible, which is supplied by the circulation pump 202 in combination.
  • Further, it is preferred to provide around the condenser a sound insulation 208 which may be active or passive. Passive sound insulation, like thermal insulation, will insulate the frequencies of the sound produced by the liquefaction process as well as possible. Likewise, it is also preferred to silencing the other components of the system.
  • The sound insulation may alternatively be actively formed, in which case z. B. would have a microphone for sound measurement and in response would trigger a sound-counteraction, such as an in-vibration displacement of an outer condenser wall, etc. with z. B. piezoelectric means.
  • This in Fig. 3a The embodiment shown is somewhat problematic in that, when the heat pump is taken out of service, the liquid 180 in the condenser penetrates into the tube 198, in which otherwise a compressed vapor is present. In one implementation, a return valve may be disposed in line 198, eg, near the exit of the line from the condenser. Alternatively, line 198 may be led up to the point where no liquid is returned to the compressor when the compressor is turned off. When the compressor is then put back into operation, so is first pressed by the compressed steam, the water from the steam line 198 in the condenser.
  • Only then is a vapor in the condenser brought to condensation when a sufficient proportion of the water has been removed from the line 198. Such a kind of embodiment thus has a certain delay time, which is needed until the water volume 180 again from the compressed Steam is warmed up. Furthermore, the work required to remove the water that has entered the conduit 198 from the conduit 198 is no longer recoverable and thus "lost" in terms of heating, such that lower efficiency losses are tolerated have to.
  • An alternative embodiment that overcomes this problem is in Fig. 3b shown. In contrast to Fig. 3a Now, the compressed steam is not supplied within a pipe below the water level in the condenser. Instead, the vapor is, so to speak, "pumped" into the liquid in the condenser from the surface. For this purpose, the condenser comprises a nozzle plate 210 which has nozzles 212 protruding with respect to the plane of the nozzle plate 210. The nozzles 212 extend below the water level of the water volume 180 in the condenser. The recessed sections between two nozzles in Fig. 3b at 214, however, extend above the water level of the water volume 180 in the condenser, so that always between two nozzles is the water surface of the condenser water, which is interrupted by a nozzle. The nozzle 212 has nozzle openings through which the compressed vapor that propagates from the conduit 198 within the vapor volume 182 may enter the condenser water, as shown schematically by arrows 216.
  • If in the implementation of Fig. 3b the compressor is taken out of service, this leads to the fact that the liquid penetrates only a little bit into the nozzles 212 of the nozzle plate 210, so that only very little work has to be spent in order to restart the operation of the heat pump, the water from the nozzles push it out again. In any event, the expander 200 ensures that, due to the guide through the expander, the liquid being pumped upwardly by the pump 202 is always as cold as possible and comes in contact with the warm vapor. The warm water then either enters immediately into the feed 20a or spreads over the Aufweiter edge in the water volume, as shown by an arrow 218, so that in the condenser outside the Aufweiters a temperature stratification occurs, which is disturbed as little as possible, in particular due to the expander shape.
  • The flow velocity at the edge of the expander, that is, where the arrow 218 is indicated, is substantially lower than in the middle. It is preferred to operate the condenser as a temperature-layer storage such that the heat pump, and in particular the compressor, does not have to run continuously, but only has to run when needed, as is the case for normal heating systems, e.g. working with an oil burner is also the case.
  • Fig. 3c shows a further preferred implementation of the condenser in a schematic form. In particular, the condenser comprises a gas separator 220, which is coupled to the gas volume 182 in the condenser. Gas generated in the condenser, such as oxygen or another gas which may outgas in the condenser, accumulates in the gas separator vessel 220. This gas can then be pumped to the atmosphere by operating a pump 222, preferably at certain intervals, since continuous gas pumping is not necessary due to the small amount of gas produced. Alternatively, the gas can also be returned to the return 112 or 113 of Fig. 2 be docked, so that the gas is brought back together with the returning groundwater back into the groundwater reservoir, where it is then dissolved in the groundwater again, or when it enters the groundwater reservoir, there goes into the atmosphere.
  • After the system of the invention works with water, even with strong outgassing no gases that were not previously solved in the groundwater, so that the separated gas has no environmental problems in itself. Again, it is emphasized that due to the turbomachinery compaction according to the invention and the use of water as the working fluid at any point contamination or contamination by a synthetic refrigerant or by an oil due to an oil circuit occur. The system according to the invention has at each point as a working medium water or Steam which is at least as clean as the original groundwater, or even cleaner than the groundwater due to evaporation in the evaporator, since it is distilled water when the compressed steam in the condenser has been re-liquefied.
  • Subsequently, reference will be made to Fig. 4a a preferred embodiment of the evaporator shown in order to use the condenser flow advantageously for accelerating the evaporation process. For this purpose, the process, which is yes to heating return temperature, that is, a much higher temperature than the groundwater funded from the ground, passed through the expander 108 of the evaporator, so that the wall of the drain pipe 204 as a germ for a Bubble boiling acts. This is much more efficient steam generated by the evaporator than if no such germination is provided. Furthermore, it is preferred to make the wall of the drainage pipe 204 as structured as possible, at least in the expander, as shown at 206, in order to further improve nucleation for the bubble boiling. The rougher the surface of the downcomer 204 is, the more nuclei are created for the bubbling. It should be noted that the passage through the drainpipe 22 is only very small, as it is only 4 ml per second added to the condenser in one operating mode. Nevertheless, even with this small amount and due to the relatively high temperature compared to the groundwater, the much more efficient bubble boiling can be brought about in order to reduce the size of the evaporator with the same efficiency of the heat pump.
  • To accelerate the evaporation process, alternatively or additionally, a region of the evaporator on which water to be evaporated is located, that is to say the surface of the expander or a part thereof, may be made of a rough material in order to supply nuclei for nucleate boiling. Alternatively or additionally, a rough grid (near) can be arranged below the water surface of the water to be evaporated.
  • Fig. 4b shows an alternative implementation of the evaporator. During the expiration in Fig. 4a has been used merely as a "pass-through" blistering aid for efficient evaporation and, as it is known in US Pat Fig. 4a left in the picture, then, when it has passed through the evaporator, is discharged, the drain in Fig. 4b itself used to enhance blistering. For this purpose, the condenser outlet 22 of Fig. 2 optionally via a pump 192 or, if circumstances permit, without a pump, connected to a nozzle tube 230 having a termination 232 at one end and nozzle orifices 234. The warm condenser water, which is discharged from the condenser via the drain 22 at a rate of, for example, 4 ml per second, is now fed to the evaporator. It will evaporate on its way to a nozzle opening 234 in the nozzle tube 230 or directly at the outlet to a nozzle due to the low pressure for the temperature of the drain water to a certain extent below the water surface of the evaporator water.
  • The resulting vapor bubbles are directly as boiling nuclei for the evaporator water, which is conveyed via the inlet 102, act. This can be triggered without major additional measures efficient bubble boiling in the evaporator, this triggering similar to Fig. 4a due to the fact that the temperature near the rough region 206 in Fig. 4a or in the vicinity of a nozzle opening 234 is already so high that immediately takes place at the present pressure evaporation. This evaporation forces the generation of a vapor bubble which, if the conditions are favorably chosen, has a very high probability that it will not collapse again, but that it will develop into a surface-going vapor bubble which, as soon as it enters the Steam volume has entered the evaporation chamber, is sucked through the suction pipe 12 from the compressor.
  • This in Fig. 4b As shown embodiment requires that condenser water is brought into the groundwater circuit, since the emerging from the nozzle tube 230 medium ultimately via the overflow of the evaporator again enters the return 112 and is thus brought into contact with the groundwater.
  • If there are any requirements under water law or any other reason that this is not permissible, then this may apply in Fig. 4c shown embodiment can be used. Here, the warm condenser water supplied from the condenser outlet 22 is introduced, for example, at a rate of 4 ml per second into a heat exchanger 236 to deliver its heat to a groundwater coming from the main groundwater flow in the line 102 via a branch line 238 and a branch pump 240 has been branched off. The branched groundwater then substantially decreases the heat of the condenser outlet within the heat exchanger 236, so that preheated groundwater, for example, at a temperature of 33 ° C is introduced into the nozzle tube 230, by the high compared to the groundwater temperature, the nucleate in Effectively trigger or support evaporators. In contrast, the heat exchanger via a drain line 238 relatively strongly cooled drain water, which is then fed via a drain pump 240 of the sewer. Due to the combination of branch line 238 and branch pump 240 and heat exchanger 236 only groundwater is used in the evaporator or introduced without it was in contact with another medium. A relevance in terms of water law thus exists in the Fig. 4c not shown embodiment.
  • Fig. 4d shows an alternative implementation of the evaporator with edge feed. Here is, as opposed to Fig. 2 , the expander 200 of the evaporator located below the water level 110 in the evaporator. As a result, water flows "from the outside" into the center of the expander and is then returned to a central conduit 112. While the center line in Fig. 2 has served to feed the evaporator, it serves in Fig. 4d now for deriving the unevaporated groundwater. In contrast, the in Fig. 2 shown line 112 for the removal of unevaporated groundwater served. In FIG. 4d On the other hand, this line on the edge acts as a groundwater supply.
  • Fig. 4e shows a preferred implementation of Aufweiters 200, as it can be used in the evaporator, or the expander, as it can be used for example in the condenser, and as it is for example in Fig. 2 or Fig. 3a or 3b is shown. The expander is preferably designed so that its small diameter preferably enters the expander at the center of the "large" expander surface. This diameter of this inlet or outlet (in Fig. 4d ) is preferably between 3 and 10 cm and in particularly preferred embodiments between 4 and 6 cm.
  • The large diameter d 2 of the expander is in preferred embodiments between 15 and 100 cm and is smaller than 25 cm in particularly preferred embodiments. The small version of the evaporator is possible when efficient measures for triggering and supporting the bubble boiling are used, as explained above. Between the small radius d 1 and the large radius d 2 there is a curvature region of the expander, which is preferably designed to give a laminar flow in this region which is of a fast flow rate, preferably in the range of 7 to 40 cm per second, is lowered to a relatively small flow rate at the edge of the expander. Strong flow rate discontinuities, such as vortices in the area of the line of curvature or "bubbling effects" above the inlet when viewed from the top of the expander, are preferably avoided as they may be detrimental to efficiency.
  • In particularly preferred embodiments, the expander has a shape which causes the height of the water level above the expander surface to be less than 15 mm and preferably between 1 and 5 mm. It is therefore preferred to use an expander 200 which is designed such that in more than 50% of the area of the expander, when viewed from above, there is a water level which is smaller than 15 mm. This ensures efficient evaporation over the entire area, which is particularly enhanced in terms of efficiency when using measures to trigger bubble boiling.
  • The heat pump according to the invention thus serves for efficient heat supply of buildings and no longer requires working equipment that has a global climate-damaging influence. According to the invention, water is evaporated under very low pressure, compressed by one or more turbomachines arranged one behind the other and liquefied again into water. The transporting energy is used for heating. According to the invention, a heat pump is used, which is preferably an open system. Open system here means that groundwater or other available thermal energy-carrying aqueous medium is evaporated, compacted and liquefied under low pressure. The water is used directly as a working medium. The contained energy is therefore not transferred to a closed system. The liquefied water is preferably used directly in the heating system and then fed back to the groundwater. To decouple the heating system capacitively, it can also be completed via a heat exchanger.
  • The efficiency and usefulness of the present invention will be illustrated by way of a numerical example. If an annual heat requirement of 30,000 kWh is assumed, according to the invention about a maximum of 3750 kWh of electric power must be expended for the operation of the turbomachine, since the turbomachine only has to supply about one eighth of the total heat requirement.
  • The eighth comes from the fact that only in the most extreme cold a sixth must be spent, and z. B. at transition temperatures such as in March or the end of October, the efficiency can rise to a value greater than 12, so that a maximum of one-eighth must be spent on average over the year.
  • With electricity costs of about 10 cents per kWh, which can be achieved for electricity when electricity is purchased, for which the power plant does not have to guarantee freedom of interruption, this corresponds to an annual cost of about 375 euros. If you want to produce 30,000 kWh of oil, you would need about 4,000 liters, which would be equivalent to a price of 2800 euros given current oil costs, which are unlikely to fall in the future. According to the invention, one can therefore save 2425 euros per year! It should also be noted that compared to the combustion of oil or gas for heating purposes by the inventive concept, up to 70% of the amount of released CO 2 is saved.
  • In order to reduce the manufacturing costs and also to reduce the maintenance and assembly costs, it is preferable to carry out the housing of the evaporator, the compressor and / or the condenser and also especially the radial impeller of the fluid machine made of plastic and in particular of injection-molded plastic. Plastic is well suited, since plastic is corrosion-resistant with respect to water and, according to the invention, advantageously the maximum temperatures are significantly below the deformation temperatures of usable plastics in comparison with conventional heaters. Furthermore, the assembly is particularly simple, since there is negative pressure in the system of evaporator, compressor and condenser. As a result, the requirements for the gaskets are significantly reduced because the entire atmospheric pressure helps to keep the housings tight. Plastic is also particularly well, since at no point high temperatures occur in the system according to the invention, which would require the use of expensive special plastics, metal or ceramic. By plastic injection molding and the shape of the radial wheel can be arbitrarily optimized and yet easily and inexpensively manufactured in spite of complicated shape.

Claims (9)

  1. A heat pump comprising:
    an evaporator device (10) comprising:
    a water evaporator (10) for evaporating water as a working liquid to generate a working vapor, the water evaporator comprising an evaporation chamber (100) and being adapted to generate an evaporation pressure of less than 20 hPa within the evaporation chamber, so that the water will evaporate at temperatures below 18°C,
    a compressor (16) coupled to the evaporator (10) for compressing the working vapor, the compressor being adapted as a continuous flow machine and further being adapted to compress the working vapor to a working pressure of more than 5 hPa above the evaporation pressure; and
    a liquefier (18) for liquefying a compressed working vapor, the liquefier being adapted to output a heat which has been acquired during the liquefaction to a heating system,
    characterized in that the water evaporator (10) further comprises
    an expander (108) which expands within the evaporation chamber to at least three times a diameter of a feed to the evaporation chamber;
    a reception apparatus (110) adapted to receive any working liquid overflowing over an edge of the expander (108); and
    a drain (112) adapted to carry off the overflowing working liquid.
  2. The heat pump as claimed in claim 1, wherein the evaporator device comprises a turbine (150), via which a pressure of an upstreaming working fluid is reduced, and which extracts energy from the working liquid in the process, the turbine (150) further being operatively coupled to a pump (152) to bring a downstreaming working liquid from the pressure present within the evaporation chamber to the pressure of the upstreaming working fluid, the operative coupling (154) being adapted such that the pump (152) uses at least part of the energy the turbine has extracted.
  3. The heat pump as claimed in claim 1, wherein the drain (112) is coupled to a flow controller (114), the flow controller (114) being controllable to maintain a level of the overflowing working liquid within the reception apparatus (110) within a predefined range.
  4. The heat pump as claimed in one of the preceding claims, wherein the evaporator device (10) comprises a gas separator adapted to remove at least part of a gas dissolved in the water to be evaporated from the water to be evaporated, so that the removed part of the gas is not sucked in by the compressor via the evaporation chamber.
  5. The heat pump as claimed in claim 4, wherein the gas separator is arranged to feed the removed part of the gas to non-evaporated water so that the gas is transported off by the non-evaporated water.
  6. The heat pump as claimed in claim 1 or 2, wherein the evaporator device (10) further comprises:
    an expander (200) which expands, within the evaporation chamber, to at least three times a diameter of the drain (112) outside the evaporation chamber;
    a reception apparatus adapted to receive the working liquid fed to the evaporation chamber; and
    an inflow to supply the reception apparatus with ground water;
    wherein the expander is arranged within the evaporation chamber such that the working fluid flows off, over an edge of the expander comprising a large diameter, to an area of the expander comprising a low diameter, and from there to a drain.
  7. The heat pump as claimed in claim 2, wherein the liquefier (18) comprises a drain (22) for draining off liquefied working liquid, the drain being coupled, at a coupling position (194), to the riser pipe (102) or to a backflow pipe (113), where a liquid pressure within the riser pipe or the backflow pipe (113) is smaller than or equal to a pressure present at the drain (22).
  8. The heat pump as claimed in claim 2 or 7, wherein the liquefier (18) comprises a drain (22) for draining off liquefied working liquid, the drain being coupled, at a coupling position (194), to the riser pipe (102) or to a backflow pipe (113), a pressure compensator (192) being arranged between the drain (22) from the liquefier (18) and the coupling position (194), the pressure compensator (192) being adapted to control a pressure of the water drained off from the liquefier (18) such that the water will enter into the riser pipe (102) or into the backflow pipe (192).
  9. A method of pumping heat, comprising:
    evaporating (10) water as a working liquid by a water evaporator (10) to generate a working vapor, the working vapor being generated at an evaporation pressure of less than 20 hPa, so that the water will evaporate at temperatures below 18°C, the water evaporator comprising
    an expander (108) which expands in the evaporation chamber to at least three times a diameter of a feed line to the evaporation chamber;
    a reception apparatus (110) adapted to receive any working liquid overflowing over an edge of the expander; and
    a drain (112) adapted to carry off the overflowing working liquid;
    compressing (16) the working vapor in terms of flow so as to compress the working vapor to a working pressure of more than 5 hPa above the evaporation pressure; and
    liquefying (18) a compressed working vapor to output a heat which has been acquired during the liquefaction to a heating system.
EP06724016A 2006-04-04 2006-04-04 Heat pump Active EP2016349B1 (en)

Priority Applications (2)

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DE202006005461U DE202006005461U1 (en) 2006-04-04 2006-04-04 Heat pump for pumping heat, has compressor implemented as turbo machine and designed to compress operating steam at operating pressure higher than five hecto-Pascal above evaporation pressure
PCT/EP2006/003061 WO2007118482A1 (en) 2006-04-04 2006-04-04 Heat pump

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
EP11158798A EP2341301A3 (en) 2006-04-04 2006-04-04 Heat pump
EP11158786.1A EP2341300B1 (en) 2006-04-04 2006-04-04 Heat pump
EP10189256.0A EP2290305B1 (en) 2006-04-04 2006-04-04 Evaporator
EP11158793.7A EP2343489B1 (en) 2006-04-04 2006-04-04 Heat pump

Related Child Applications (7)

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EP10189256.0A Division EP2290305B1 (en) 2006-04-04 2006-04-04 Evaporator
EP11158786.1A Division EP2341300B1 (en) 2006-04-04 2006-04-04 Heat pump
EP11158793.7A Division EP2343489B1 (en) 2006-04-04 2006-04-04 Heat pump
EP10189256.0 Division-Into 2010-10-28
EP11158798.6 Division-Into 2011-03-18
EP11158793.7 Division-Into 2011-03-18
EP11158786.1 Division-Into 2011-03-18

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EP2016349A1 EP2016349A1 (en) 2009-01-21
EP2016349B1 true EP2016349B1 (en) 2011-05-04

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EP10189256.0A Active EP2290305B1 (en) 2006-04-04 2006-04-04 Evaporator
EP11158798A Withdrawn EP2341301A3 (en) 2006-04-04 2006-04-04 Heat pump
EP11158786.1A Active EP2341300B1 (en) 2006-04-04 2006-04-04 Heat pump
EP06724016A Active EP2016349B1 (en) 2006-04-04 2006-04-04 Heat pump

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EP10189256.0A Active EP2290305B1 (en) 2006-04-04 2006-04-04 Evaporator
EP11158798A Withdrawn EP2341301A3 (en) 2006-04-04 2006-04-04 Heat pump
EP11158786.1A Active EP2341300B1 (en) 2006-04-04 2006-04-04 Heat pump

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US (3) US7841201B2 (en)
EP (4) EP2290305B1 (en)
JP (1) JP5216759B2 (en)
DE (2) DE202006005461U1 (en)
WO (1) WO2007118482A1 (en)

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US20160109139A1 (en) 2016-04-21
JP2009532655A (en) 2009-09-10
WO2007118482A1 (en) 2007-10-25
EP2341300B1 (en) 2017-09-06
DE502006009456D1 (en) 2011-06-16
JP5216759B2 (en) 2013-06-19
US20110036100A1 (en) 2011-02-17
US20070245759A1 (en) 2007-10-25
US7841201B2 (en) 2010-11-30
US9222483B2 (en) 2015-12-29
EP2290305A1 (en) 2011-03-02
US10337746B2 (en) 2019-07-02
DE202006005461U1 (en) 2007-08-16
EP2341301A2 (en) 2011-07-06
EP2341300A1 (en) 2011-07-06

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